Effect of Tip Clearance on Performance of a Centrifugal Compressor

IJRET: International Journal of Research in Engineering and Technology eISSN: 2319-1163 | pISSN: 2321-7308
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Volume: 02 Issue: 09 | Sep-2013, Available @ http://www.ijret.org 445
EFFECT OF TIP CLEARANCE ON PERFORMANCE OF A
CENTRIFUGAL COMPRESSOR
S.M.Swamy1
, V.Pandurangadu2
1
Assistant Professor, Dept .of Mechanical Engg, GNITS, Shaikpet, Hyderabad, India -5000 08, moodswamy@yahoo.co.in
2
Professor, Dept of Mechanical Engg, JNTUA, Ananthapur, India
Abstract
The centrifugal compressor is to study the effect of tip clearance on the performance characteristics and the wall static pressure for a
different flow co-efficient. The method of testing the compressor is run at a constant speed at 2000rpm. The tip clearance is varied by
using spacers. The volume flow rate is varied with the help of throttling device to conduct the performance test. The performance
characteristic of the centrifugal compressor showing the variation of discharge pressure with volume flow rate is plotted. Obtaining
the performance characteristics showing the variation of discharge pressure with volume flow rate for different tip clearance, viz.
=2.2%, 4%, 6.1% and 7.9%. Measurement of periodic pressure at various tip clearance viz. =2.2%, 4%, 6.1% and 7.9%. For
each tip clearance pressure measured in radial location of impeller at 6 positions for different flow co efficient values. Five flow
coefficients viz., ф =0.40, =0.34 (both above design flow), =0.28 (near design flow), =0.21=0.18 (both below design flow) and four
values of non-dimensional tip clearance viz., =2.2%, 4%, 6.1% and 7.9%, are chosen for experimental work. The objective of the
research work is to measure the periodic variation static pressure on the casing over the rotor at different values of tip clearance and
flow coefficients. With the availability of these data, it is possible to improve the tip clearance flow in centrifugal compressor.
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1. INTRODUCTION
Tip clearance in centrifugal compressor causes the leakage of
high pressure fluid from pressure surface to suction surface of
the impeller blade, making the flow field highly complex and
affecting the performance. Hayami (1997) has found from his
experiments that axial movement of the casing has better
efficiency over the movement of casing in radial and axial
directions. Radial movement of casing increases clearance at
inducer, which reduces the operating range. The tip clearance
studies are conducted to understand the flow behavior in order
to minimise the effect of tip clearance. Pampreen (1973),
Mashimo et al. (1979), Sitaram and Pandey (1990) have
conducted experimental studies and suggested that by
reducing the tip clearance gap size, the tip clearance effect can
be minimised. The effect of tip leakage on flow behavior in
rotating impeller passage was carried out by Hathaway et al.
(1993). Farge et al. (1989) carried out the impeller passage
measurements at five stations in a low speed centrifugal
compressor using a five hole probe rotating with the impeller.
The flow in the impeller is affected by many geometric and
flow parameters. Due to its extreme complexity, the
characteristics of the flow inside a rotating impeller are the
least understood phenomenon in the turbo machinery aero
dynamics. In centrifugal compressor, there are mainly of two
types of impellers shrouded and unshrouded. In shrouded
impeller, the shroud rotates with the impeller and so, there is
no clearance between the shroud and the tip of the blades. In
case of unshrouded impellers the front cover of the casing
forms the stationary shroud and so, there is a narrow clearance
between the tip of the blade and the shroud. While running,
there is a leakage flow, through this tip clearance which
deteriorates the performance of the machine. i.e. reduces the
pressure ratio, increases the input power for decreases the
efficiency. Now, the question is why one should go for an
unshrouded impeller, while there is no tip clearance in case of
a shrouded impeller. The reason being that in case of shrouded
impellers, there are secondary flows from the pressure side of
the blade to the suction side of the adjacent blade along the
shroud. Over the last two decades or so there has been a
requirement to measure time-varying pressures in turbo
machinery applications to bandwidths of order 100 kHz and
the silicon piezo resistive pressure sensor has been the device,
which has been at the heart of many of the measurements. In
case of unshrouded impellers the leakage flow through the tip
clearance of blades deteriorates the performance of the
machine. At the exit, since the specific volume of the gas is
reduced considerably (pressure being high).i.e. tip clearance
effect is more significant. The impeller geometry should be
designed considering tip clearance effects. The leakage flow
through tip clearance of the impeller blades also modifies the
flow pattern in the impellers. Many investors have examined
the details of the flow pattern, but the results have not been
utilized to evaluate the pressure loss due to tip clearance of
turbomachine.
2. EXPERIMENTAL INVESTIGATIONS
The effect of tip clearance on the performance of a centrifugal
compressor has been studied by many persons. Ishida and

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Senoo (1981) measured pressure distribution along the shroud
at seven different tip clearances. They concluded that the
pressure loss due to tip clearance is proportional to the
pressure rise due to deceleration of relative velocity. They also
observed that the change in input power due to a change of tip
clearance is related to the effective blockage at the impeller
exit. Ishida and Senoo (1986) modified their own theory on tip
clearance loss of centrifugal impellers as Efficiency drop due
to tip clearance of high pressure ratio compressors is less than
that of low pressure ratio compressors, if the tip clearance
ratios at the impeller exit in equal. The magnitude of clearance
loss becomes smaller as the flow rate is reduced and also at a
reduced shaft speed in cases of high pressure ratio
compressors. Freeman (1985) suggested that 1% increase in
tip clearance produced by increasing the casing diameter will
give approximately 1.4% loss in efficiency Schumann et al.
(1987) tested a centrifugal impeller with a vane less diffuser
for four impeller area ratios from 2.322 to 2.345. The impeller
was initially designed for a pressure ratio of approximately 5.5
and a mass flow rate of 0.959Kg/s. For each area ratio a series
of impeller exit axial clearances was also tested. They found
that impeller efficiency decreases by 0.4 point for every 1%
increase in exit clearance. Wisler (1985) reported that for a
low speed four stage compressor the effect of increasing rotor
tip clearance from 1.6% to 3.4% was to reduce the efficiency
by 1.5% and the peak pressure rise by 9.7% Smith and
Cumpsty (1984) have shown a 23 percent drop in maximum
pressure rise and a 15 percent increase in flow coefficient at
stall in a large, low speed compressor as the tip clearance was
increased from 1 to 6 percent of chord. Engeda and
Rautenberg (1987) carried out experiments on five centrifugal
pump impellers of different specific speeds ranging from 0.33
to 1.51 and studied the relative effect on tip clearance on the
stage performance. They concluded that the tangential
component of the absolute velocity is seen to be hardly
influenced by the tip clearance variation. Leakage losses and
so the overall efficiency decrement increased with increase in
tip clearance due to increase in flow passage. Also, a clearer
dependency of clearance effects on specific speed was not
observed. They concluded that the problems tip losses could
most likely be well understood from detailed impeller
geometry, impeller relative velocity distribution measurements
for secondary flow analysis.
3. COMPUTATIONAL STUDIES
With the advances in memory and speed of computers, the
inviscid flow analysis for turbo machinery has progressed
from the two-dimensional to the quasi-three dimensional flow
analysis; computer time required is very large for it is not
practical to use this analysis for the design of turbo machine.
To reduce computer time, it is desirable to reduce the number
of grid points in each cross section. A flow analysis between
blades with the tip clearance requires many grid points
because the grid must be considerably smaller than the
clearance while the leakage flow rolls up, forming a vortex of
a diameter about ten times of the tip clearance. Moore et al.
(1984) developed a partially parabolic calculation procedure to
calculate three dimensional viscous flows in a centrifugal
impeller. The three dimensional pressure fields within the
impeller was obtained by first performing a three dimensional
inviscid flow calculation for then by adding a viscosity model
and a viscous wall boundary condition to allow calculation of
the three-dimensional viscous flow. The basic idea is that the
clearance velocity field can be approximately decomposed in
to independent through flow and cross flow, since chord wise
pressure gradients are much smaller than normal pressure
gradients in the clearance region. As in the slender body
approximation in external aerodynamics, this description
implies that the three dimensional, steady, clearance flows can
be viewed as a two dimensional, unsteady flow. Using this
approach, a similarity scaling for the cross flow in the
clearance region is developed and a generalized description of
the clearance vortex is derived. The predicted results are
compared with the experimental results measures at the exit of
the impeller using a two-hole probe with high-frequency
response semiconductor pressure transducer developed by
Goto (1988). Predicted internal flow, the interaction
mechanism between passage vortices and tip leakage flow and
the effects of increasing tip clearance were discussed.
4. EXPERIMENTAL FACILITY AND
INSTRUMENTATION
The experimental set up of centrifugal compressor; driven by
5KW motor. The main components of the compressor are the
nozzle at the inlet, the suction duct, the impeller, the vane less
diffuser formed by the front and the rear walls of the casing
and volute casing of circular cross section and a delivery duct
with a throttle at its outlet. The D. C. Motor is directly coupled
to the shaft carrying the impeller. A brief description of each
of the component of the experimental set-up is given below.
An inlet nozzle and a suction duct were provided at the inlet of
the compressor. The purpose of the nozzle at the inlet of the
suction duct is to provide uniform flow to the impeller by
accelerating it. Besides, the inlet nozzle is used to measure the
volume flow through the impeller. An inlet duct of 300 mm
diameter and 1300 mm long is provided at the inlet to the
centrifugal compressor. Since the inlet duct is long, the flow
will settle by the time it approaches the impeller inlet. The
front cover of the casing was designed such a way that its
inner surface was similar to the shape of impeller blade shroud
contour for its extension formed the front wall of the vaneless
diffuser at the exit of the impeller. Six static pressure holes
were made along the meriodional plane on the inner contour
surface of the shroud from the inlet end to the exit end of the
diffuser to measure the static pressures. All the tapping was
normal to the inner contour of the shroud. An impeller of a
centrifugal compressor is the most delicate part to be
designed. It is to be ensured that the flow is ideally oriented in
the channel passages. The energy transfer from the shaft to the
fluid takes place in the rotating impeller. The existing impeller

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[6] (Sankaran, 1986) is a special kind, designed to consider the
effect of the outlet edges of the vanes of a centrifugal impeller
on its performance and flow characteristics. The impeller was
designed on the power and speed of the available drive motor.
The shape number was chosen was 0.1. The impeller is
unshrouded one, with an inducer at the entry. The inducer is
provided to give a smooth guidance to the flow turning from
axial to radial direction inside the inlet region of the impeller.
The ratio of the inducer tip diameter to the impeller outer
diameter, d1t/d2t, is 0.6 to minimize the frictional losses. The
blade inlet angle β2 at the impeller tip is 350. The meridional
component of velocity is uniform at the inlet of the impeller. It
is nearly constant from the inlet to the exit of the impeller. The
number of blades used in the impeller is 16. The selection of
the number of blades is based on Pfleiderer’s equation, which
gives the required number corresponding to the impeller
geometry. If an excessive number of blades are used, lesser
frictional and blockage would be high reducing the efficiency.
Also, if lesser number of blades were used there would not
have been proper guidance of the fluid in the impeller
resulting in high slip and lowering of the work addition to the
fluid. With the lesser number of blades, the blade loading
would also be higher. The impeller has twisted blades at the
exit region. The impeller blade edges have a maximum twist
angle of 150 at the hub and shroud ends. This makes the blade
angles at the exit of the impeller to uniformly vary from 750
from the hub end to 900 at the mid span and then to 1050 at
the shroud end. All angles are measured with respect to the
tangential direction. The rear cover of the casing called as
back shroud forms the hub wall and is attached to the volute
casing by means of studs. A minimum clearance between the
impeller and the rear cover is provided to minimize the
leakage. The cover of the casing is extended to form the back
wall of the vane less diffuser. The vaneless diffuser is a simple
device consisting of an annulus in which the radial velocity
component is reduced by an increase in the area and the
tangential velocity by the requirement of constant fluid
angular momentum (free vortex) .The extended part of the
parallel walls of front cover and the rear cover formed
vaneless diffuser. The width of vaneless diffuser was kept
same to direct the flow from, impeller to the annular passage
without lateral expansion. The function of the volute casing is
to collect the air from the vane less diffuser and discharges it
to the delivery duct. The volute casing was of circular cross
section. The width at entry to the volute casing is also kept the
same as that of the vane less diffuser exit width, so that the
surfaces at the two sides of the volute at entry forms a
continuous portion of the vaneless diffuser. It is firmly fixed to
the foundation. The volute casing is designed for a constant
angular momentum for the air stream. The cross-section of the
volute is circular. It has a base circle of 630 mm diameter, a
volute angle of 13o and cut-off angle of 23o. The passage
width at the entry to the volute cross-section at the exit is 200
mm in diameter and is directly connected to a delivery duct of
the same diameter through a 90o bend. The delivery duct is
200 mm in diameter and 5 m in length. The throttle cone is
fitted at the end of the delivery duct to regulate the volume
flow through the impeller. The average delivery pressure is
measured through the static pressure tapping provided on
delivery duct wall at a distance of about 1 m from the 90o
bend following the volute casing. The throttle cone could be
moved forward or backward manually inside the delivery duct
to regulate the volume flow through the compressor. On the
shroud adapters, for the pressure sensor, are provided at six
different radial locations from inlet, just behind the leading
edge, to exit, after the trailing edge of the impeller (Fig. 1).
The sensor is passed inside a brass holder of 2 mm internal
diameter and with external threading (M4). The sensor is
glued to this brass holder using Fevikwik. The brass holder is
screwed into these adapters at different locations for pressure
measurement. Spacers were used to maintain the tip clearance
accurately. Tip clearance is varied by moving the front shroud
axially. Thereby the distance between the front and rear cover
would alter. Provisions are made on front cover to fix the
spacers in position. Four spacers were used for each tip
clearance were fixed at 90 degrees. The experimental
investigations involve study for four tip clearance .so hence
four set of spacers were made. To measure the suction,
delivery and wall static pressures along the shroud and
scanning box (Furnace Control Model no. FCO 91-3) with a
micro manometer (Model FCO12) manufactured by M/s
Furness Control Ltd., Bexhill, UK is used. The micro
manometer has a sensitive differential pressure measuring
device capable of reading air pressures up to ± 1999 mm WG.
It would respond to pressure inputs up to 5 kHz. But the time
constant potentiometer could be used to dampen response the
fluctuating signal. It is useful in damping flow fluctuations.
The pressure can be measured to an accuracy of ± 0.1mmWG.
The scanning box contains 20 valves so we can measure up to
20 pressures. The speed of the centrifugal compressor is
measured using a non-contact digital type tachometer. The
speed measurement is based upon the evaluation of reflecting
pulses from the rotating shaft. It has a range of 0-9999 and its
accuracy is ± 1 rpm.
Fig. 1 Meridional view of impeller showing static pressure
measurement

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Table 1 Design Details of the Rotor
Total pressure rise, ∆p: 300 mm WG
Volume flow rate, V: 1.12 m3/s
Speed of rotation, N: 2000 rpm
Shape number, Nsh: 0.092
No. of rotor blades, Z: 16
Inducer hub diameter, d1h: 160 mm
Inducer tip diameter, d1t: 300 mm
Rotor tip diameter, d2: 500 mm
Blade height at the exit, h2: 34.74 mm
Blade angle at inducer tip,
β1t:
35°
Blade angle inducer hub,
β1h:
53°
Blade angle at exit, β2: (a) At hub: 75°
(b) At mean section: 90° (c) At tip:
105°
All the angles are measured w. r. t. tangential
direction
Fig. 2 Details of Configurations Tested
The non-dimensional parameters are defined as follows:
Flow coefficient, φφφφ =
22 Ubdπ
r2
22
c
U
V
=
Where V = Volume flow (m3
/s)
d2 = Rotor tip diameter (m)
b2 = Rotor blade width at exit (m)
c2r = Radial velocity at rotor exit (m/s)
U2 = Rotor tip speed (m/s)
Energy coefficient, ψψψψ =
2
2
U
W2
Where W = Specific work (m2
/s2
)
= ΔZ
ρ
sp
d
p
g
cc
2
22
sd +
−
+
−
Power coefficient γγγγ =
3
2
3
2
UA
EIm2
UA
N2
ρ
η
ρ
c =
where Nc = Coupling power (Watts)
A = Suction duct area (m2
)
E = Motor voltage (Volts)
I = Motor current (Amps)
ηm = Motor efficiency
Compressor efficiency, ηηηηc =
c
ρ
N
WV
Table 1.1 Colour designations of sensor wires
5. RESULTS AND DISCUSSION
The results of experiments conducted on the centrifugal
impeller at four values of tip clearances i.e., =2.2%, 4%, 6.1%
and 7.9% for f0ur different flow coefficients i.e., = 0.18, 0.21,
0.28, 0.34 and 0.40 are presented and discussed in this chapter.
The performance characteristics of the impeller are presented
first and then results of wall static pressure measured from the
front shroud to impeller exit are presented.
6. PERFORMANCE CHARACTERESTICS
Performance test is carried out at a constant speed at varying
throttling positions from low flow rate (in surge region) to
high volume flow rate (near choking). Five flow coefficients
viz., =0.40, =0.34 (both above design flow), =0.28 (near
design flow), =0.21=0.18 (both below design flow) and four
values of non-dimensional tip clearance viz., =2.2%, 4%,
6.1% and 7.9% are chosen for experimental work. The
performance of the compressor in terms of energy coefficient
COLOUR DESIGNATION
RED + INPUT TO SENSOR
BLACK - INPUT TO SENSOR
GREEN + OUTPUT FROM
WHITE - OUTPUT FROM

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vs. flow coefficient is presented in Fig. 5.1. The range of flow
coefficient for which the compressor operates stably is =0.08
to =0.48.The result of the experimental investigations revealed
that, energy coefficient across the compressor is reduced as
the tip clearance is increased. The flow coefficient, where the
energy coefficient is maximum decreases with the increase in
tip clearance. Similar results are observed in a low speed
radial tipped centrifugal compressor (Sitaram and Pandey,
1990). However the operating range of the compressor is
decreased slightly as the maximum volume flow passing
through the compressor is reduced with tip clearance. The
performance curves shown include losses not only the in the
impeller, but also in the following vaneless diffuser, volute
casing and delivery duct. The static pressure distribution
measured at the casing static pressure holes is shown in Fig.
5.2, at 2.2% tip clearance for different flow coefficients. On
the shroud, static pressure taps were provided from the inducer
leading edge to the impeller exit. The shroud has its inner
shape contoured to match the impeller blade tip profile from
the inlet to the exit. The static pressure measurements on the
shroud indicate variation in the static pressure coefficient,
which matches with the blade passage clearly. From the figure
it is seen that the distribution is high for ф= 0.28, as it is
nearer to design point operation. The pressure initially
decreases due to suction and then uniformly increases,
indicating that there are no dead zones or eddies near the
shroud region inside the impeller and energy transfer occurs
smoothly to the fluid near the shroud. This can be inferred
from the Fig 5.1. It is lesser in the case of other flow
coefficients. The first two static pressure tapping shows a
negative static pressure coefficient. Very close to the third
hole position the leading edge of the impeller starts. Since
then, the pressure coefficient took an increasing trend. This is
due to the development of low pressure zone behind the blade
in the suction side.
The compressor characteristic curve against flow coefficient at
different tip clearances is shown in Fig. 5.3. Using the micro
manometer the static pressure difference between the entry
and exit of the compressor is measured, at a particular tip
clearance. The figure depicts a negligible difference in
efficiency for flow coefficients up to 0.30, for all tip
clearances. At flow rates higher than this, the difference in
efficiency is appreciable, the peak efficiency being lower for
higher tip clearance and vice versa. This can be attributed to
the increase in tip losses with the increase in the clearance.
The reduction in efficiency is due to the reduction in energy
transfer for higher tip clearances, since the input power
remains more or less the same.
Figure 5.4 shows the effect of volume flow rate on pressure
ratio P2/P1 across the compressor. Pressure ratio shows an
inverse variation with increase of volume flow rate, for all tip
clearances. As the tip clearance increases, the space available
for the flow to expand is comparatively more than a lower tip
clearance; the pressure drops more and is clearly seen in the
figure. Figure 5.5 shows the effect of volume flow rate on
pressure rise (P2-P1). From the figure it is clearly seen that
when volume flow rate increases the pressure difference is
reduced. For the same tip clearance, as the volume flow rate is
increased the pressure at the exit increases for hence the
pressure difference is having a decreasing trend.
CONCLUSIONS
The present experimental investigations were conducted to
study the effect of tip clearance on the performance and wall
static pressure in a centrifugal compressor. Performance tests
were conducted at five flow coefficient with four different tip
clearances for each flow coefficient. Based on the
experimental investigations, following conclusions are drawn
1 As the tip clearance is increased, there is a reduction in
static pressure rise across the compressor, which causes
reduction in energy coefficient.
2 The decrease in energy coefficient is more at higher value
of flow coefficient.
3 The operating range slightly decreases with the increase
in tip clearance.
The wall static pressure is high for lower tip clearance for all
the flow coefficients. For high flow coefficient, reduction in
static pressure is more near the inducer, which indicates more
deceleration of flow at higher flow coefficient. The tip
clearance is more pronounced at higher flow coefficients.
From these graphs it shows that the maximum pressure is
achieved for 2.2% tip clearance, as compared to other tip
clearance, for all flow coefficients. Similarly for a flow
coefficient of 0.28, the static pressure observed is higher than
other flow coefficients for all tip clearances.
NOMENCLATURE
b : Distance between the shroud and hub at the rotor exit (m)
Cd: Velocity in delivery duct (m/s)
Cs: Velocityinsuctionduct(m/s)
d: Rotor diameter (m)
N: Rotational speed of rotor (rpm)
Nc: Coupling power (Watt)
Nsh: Shapenumber=N√V/W3/4
Pd: Delivery pressure (Pa)
Ps: Suction pressure (Pa)
t : Rotor blade clearance (m)
U: Rotor tip speed = (πdn/60) (m/s)
V: Volumeflowrate(m3/s)
W: Specific work (m2/s2)
φ: Flowcoefficient (defined in the text)
γ: Powercoefficient (defined in the text)
η: Efficiency (defined in the text)
ψ: Energycoefficient (defined in the text)

__________________________________________________________________________________________
ρ: Densityofair(kg/m3)
τ: Tip clearance as a percentage of rotor
blade
height at exit = (t/b2) × 100.
Subscript : Tip.
ACKNOWLEDGMENTS
The authors would like to thank the faculty, technical staff and
administrative staff, for their help and encouragement during
the course of the present investigation.
C
u
Tangential velocity
m Non-dimensional meridional distance
P Static pressure
P
atm
Atmospheric pressure
P
o
Total pressure
R Non-dimensional radius
u
2
Blade tip speed
x Non-dimensional axial distance
φ Flow coefficient
ρ Density
τ Tip clearance
ψ
s
Static pressure coefficient
ψ
o
Total pressure coefficient
REFERENCES
1. Engeda A. and M. Rautenberg (1987) Comparisons of the
Relative effect of Tip Clearance on centrifugal impellers,
ASME paper 87-GT-11.
2. Sieverding, C. H., Arts, T., Denos, R. and Brouckaert, J. F.,
(2000) Measurement techniques for unsteady flows in
turbomachines, Experiments in Fluids, Vol. 28, pp. 285-321.
3. Sridhara, T. N., (1998) Experimental Investigation on Tip
Clearance Effects in a Centrifugal Compressor, Ph. D. thesis,
IIT Madras
4. Ishida, M., Ueki, H. and Senoo, Y., (1990) Effect of blade tip
configuration on tip clearance loss of a centrifugal impeller,
ASME Jl. of Turbomachinery, Vol. 112, No. 1, 14-18.
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Gyarmathy, G.,
(2000) Time-resolved flow measurements with fast-response
aerodynamic probes in turbomachines, Measurement Science
and Technology, Vol. 11, 1036–1054.
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tip clearance of centrifugal blowers, Trans.ASME, Jl of
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area ratio on the performance of a 5.5:1 pressure centrifugal
impeller, Trans. ASME Jl. Of Turbomachinery, 109, 10-19.
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11 Coleman, H.W. and W.G.Steele,
Jr., (1989) Experimentation and Uncertainty Analysis for
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12 Pinarbasi, A. and M. W. Johnson
(1994) Detailed Flow Measurements in a Centrifugal
Compressor Vaneless Diffuser, Trans. ASME, Jl. of
Turbomachinery, 116, 602-610.
13 Storer, J. A. and Cumpsty, N.A,
(1991) Tip Leakage Flow in Axial Compressors. Trans. of
ASME, Jl. of Turbomachinary, 113, 252-259.
14 Sturge, D. P. and N. A. Cumpsty, (1975) Two Dimensional
Method for Calculating Separated Flow in a Centrifugal
Impeller, Trans. ASME, Jl. of Fluids Engg, 97, 581-597.
15 Senoo, Y. And M. Ishida, (1975)
Behaviour of Severely Asymmetric Flow in a Vaneless
Diffuser, Trans. ASME, Jl. of Engg for Power, 97, 375-382.
16 Senoo, Y. and M. Ishida, (1987)
Deterioration of Compressor Performance due to Tip Clearance
of Centrifugal Impellers, Trans. ASME, Jl. of Turbomachinery,
109, 55-61.
17 Senoo, Y., (1991) Mechanics on the Tip Clearance Loss of
Impeller Blades, Trans. ASME, Jl. of Turbomachinery, 113, 581-597.
18 Thomson D. W., P. I. King and D. C.Rabe, (1998) Experimental
Investigation of Stepped Tip Gap Effects on the Performance of
a Transonic Axial-Flow Compressor Rotor, Trans. of ASME, Jl.
of Turbomachinary, 120, 477-486.
19 Wood, J. R., P. W. Adam and A. E. Buggele, (1983) NASA
Low-Speed Centrifugal Compressor for Fundamental Research,
NASA TM 83398.

__________________________________________________________________________________________
Fig. 5.1 Performance characteristics
Fig. 5.2 Variation of time averaged static pressure along the casing

__________________________________________________________________________________________
Fig. 5.3 Compressor characteristic curve
Fig. 5.4 Effect of volume flow rate on pressure ratio

__________________________________________________________________________________________
Fig. 5.5 Pressure difference across the compressor

Effect of Tip Clearance on Performance of a Centrifugal Compressor

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